Various systems can be used to actively control the timing and lift of engine valves to achieve improvements in engine performance, fuel economy, emissions, and other characteristics. Depending on the means of the control or the actuator, these systems can be classified as mechanical, electrohydraulic, and electro-mechanical (sometimes called electromagnetic). Depending on the extent of the control, they can be classified as variable valve-lift and timing, variable valve-timing, and variable valve-lift. They can also be classified as cam-based or indirect acting and camless or direct acting.
In the case of a cam-based system, the traditional engine cam system is kept and modified somewhat to indirectly adjust valve timing and/or lift. In a camless system, the traditional engine cam system is completely replaced with electrohydraulic or electro-mechanical actuators that directly drive individual engine valves. All current production variable valve systems are cam-based, although camless systems will offer broader controllability, such as cylinder and valve deactivation, and thus better fuel economy.
Problems with an electromechanical camless system include difficulty associated with soft-landing, high electrical power demand, inability or difficulty to control lift, and limited ability to deal with high and/or varying cylinder air pressure. An electrohydraulic camless system can generally overcome such problems, but it does have its own problems such as performance at high engine speeds and design or control complexity, resulting from the conflict between the response time and flow capability. To operate at up to 6,000 to 7,000 rpm, an actuator has to first accelerate and then decelerate an engine valve over a range of 8 mm within a period of 2.5 to 3 milliseconds. The engine valve has to travel at a peak speed of about 5 m/s. These requirements have stretched the limit of conventional electrohydraulic technologies.
One way to overcome this performance limit is to incorporate, in an electrohydraulic system like in an electromechanical system, a pair of opposing springs which work with the moving mass of the system to create a spring-mass resonance or pendulum system. In the quiescent state, the opposing springs center an engine valve between its end positions, i.e., the open and closed positions. To keep the engine valve at one end position, the system has to have some latch mechanism to fight the net returning force from the spring pair, which accumulates potential energy at either of the two ends. When traveling from one end position to the other, the engine valve is first driven and accelerated by the spring returning force, powered by the spring-stored potential energy, until the mid of the stroke where it reaches its maximum speed and possesses the associated kinetic energy; and it then keeps moving forward fighting against the spring returning force, powered by the kinetic energy, until the other end, where its speed drops to zero, and the associated kinetic energy is converted to the spring-stored potential energy.
With its well known working principle, this spring-mass system by itself is very efficient in energy conversion and reliable. Much of the technical development has been to design an effective and reliable latch-release mechanism which can hold the engine valve to its open or closed position, release it as desired, add additional energy to compensate for frictions and highly variable engine cylinder air pressure, and damp out extra energy before its landing on the other end. As discussed above, there have been difficulties associated with electromechanical or electromagnetic latch-release devices. There has also been effort in the development of electrohydraulic latch-release devices.
Disclosed in U.S. Pat. No. 4,930,464, assigned to DaimlerChrysler, is an electrohydraulic actuator including a double-ended rod cylinder, a pair of opposing springs that tends to center the piston in the middle of the cylinder, and a bypass that short-circuits the two chambers of the cylinder over a large portion of the stroke where the hydraulic cylinder does not waste energy. When the engine valve is at the closed position, the bypass is not in effect, the piston divides the cylinder into a larger open-side chamber and a smaller closed-side chamber, and the engine valve can be latched when the open-side and closed-side chambers are exposed to high and low pressure sources, respectively, because of the resulting differential pressure force on the piston in opposite to the returning spring force. When the engine valve is at the open position, the piston divides the cylinder into a larger closed-side chamber and a smaller open-side chamber, and the engine valve can be latched by exposing a larger closed-side chamber and smaller open-side chamber with high and low pressure sources, respectively.
At either open or closed position, the engine valve is unlatched by briefly opening a 2-way trigger valve to release the pressure in the larger chamber and thus eliminate the differential pressure force on the piston, triggering the pendulum dynamics of the spring-mass system. The 2-way valve has to be closed very quickly again, before the stroke is over, so that the larger chamber pressure can be raised soon enough to latch the piston and thus the engine valve at its new end position. This configuration also has a 2-way boost valve to introduce extra driving force on the top end surface of the valve stem during the opening stroke.
The system just described has several potential problems. The 2-way trigger valve has to be opened and closed in a timely manner within a very short time period, no more than 3 ms. The 2-way boost valve is driven by differential pressure inside the two cylinder chambers, or stroke spaces as the inventers refer as, and there is potentially too much time delay and hydraulic transient waves between the boost valve and cylinder chambers. Near the end of each stroke, the larger cylinder chamber has to be back-filled by the fluid fed through a restrictor, which demands a fairly decent opening size on the part of the restrictor. On the other hand, at the onset of the each stroke, the 2-way trigger valve has to relieve the larger chamber which is in fluid communication with the high pressure fluid source through the same restrictor. During a closing stroke, there is no effective means to add additional hydraulic energy until near the very end of the stroke, which may be a problem if there are too much frictional losses. Also, this invention does not have means to adjust its lift.
DaimlerChrysler has also been assigned U.S. Pat. Nos. 5,595,148, 5,765,515, 5,809,950, 6,167,853, 6,491,007, and 6,601,552, which disclose improvements to the teachings of U.S. Pat. No. 4,930,464. The subject matter up to U.S. Pat. No. 6,167,853 resulted in various hydraulic spring means to add additional hydraulic energy at the beginning of the opening stroke to overcome engine cylinder air pressure force. One drawback of the hydraulic spring is its rapid pressure drop once the engine valve movement starts.
In U.S. Pat. No. 6,601,552, a pressure control means is provided to maintain a constant pressure in the hydraulic spring means over a variable portion of the valve lift, which however demands that the switch valve be turned between two positions within a very short period time, say 1 millisecond. The system again contains two compression springs: a first and second springs tend to drive the engine valve assembly to the closed and open positions, respectively. The hydraulic spring means is physically in serial with the second compression spring. During a substantial portion of an opening stroke, it is attempted to maintain the pressure in the hydraulic spring despite of the valve movement and thus provide additional driving force to overcome the engine cylinder air pressure and other friction, resulting in a net fluid volume increase in the hydraulic spring means and an effective preload increase in the second compression spring because of a force balance between the hydraulic and compression springs. In the following valve closing stroke, the engine valve may not be pushed all the way to a full closing because of higher resistance from the second compression spring.
A concern common to this entire family of inventions is that there have to be two switchover actions of the control valve for each opening or closing stroke. Another common issue is the length of the actuator with the two compression springs separated by a hydraulic spring. When the springs are aligned on the same axis, as disclosed in U.S. Pat. No. 5,809,950, the total height may be excessive. In the remaining patents of this family, the springs are not aligned on a straight axis, but are instead bent at the hydraulic spring, and the fluid inertia, frictional losses, and transient hydraulic waves and delays may become serious problems. Another common problem is that the closing stroke is driven by the spring pendulum energy only, and an existence of substantial frictional losses may pose a serious threat to the normal operation. As to the unlatching or release mechanism, some embodiments use a 3-way trigger valve to briefly pressurize the smaller chamber of the cylinder to equalize the pressure on both surfaces of the piston and reduce the differential pressure force on the piston from a favorable latching force to zero. Still the trigger valve has to perform two actions within a very short period of time.
U.S. Pat. No. 5,248,123 discloses another electrohydraulic actuator including a double-ended rod cylinder, a pair of opposing springs that tends to center the piston in the middle of the cylinder, and a bypass that short-circuits the two chambers of the cylinder over a large portion of the stroke where the hydraulic cylinder does not waste energy. Much like the referenced DaimlerChrysler patents, it has the larger chamber of the hydraulic cylinder connected to the high pressure supply all the time. Different from DaimlerChrysler, however, it uses a 5-way 2-position valve to initiate the valve switch and requires only one valve action per stroke. The valve has five external hydraulic lines: a low-pressure source line, a high-pressure source line, a constant high-pressure output line, and two other output lines that have opposite and switchable pressure values. The constant high pressure output line is connected with the larger chamber of the cylinder. The two other output lines are connected to the two ends of the cylinder and are selectively in communication with the smaller chamber of the cylinder. Much like the DaimlerChrysler disclosures, it has no effective means to add hydraulic energy at the beginning of a stroke to compensate for the engine cylinder air force and friction losses. It is not capable of adjusting valve lift either.
A key disadvantage of the prior-art actuators identified above will be discussed in conjunction with FIG. 21, which depicts time histories of five key pressure values at the beginning of an engine valve opening event. Note that the values at the beginning of an engine valve closing event follow the same patterns but with opposite polarities and lead to the same drawbacks. In these figures, the terms first and second fluid spaces refer to fluid volumes at the engine valve closing direction (or first direction) and opening direction (or second direction) sides of the piston, respectively, and their pressures thus tend to drive the engine valve in the valve opening direction (second direction) and closing direction (first direction). The differential pressure is equal to the difference between the first and second fluid space pressure values, which assists and impedes the engine valve opening when positive and negative, respectively. The first and second ports feed fluid to the first and second fluid spaces, respectively, and the ports themselves are designed so that they are switched in predetermined ways to complete their latch and release functions.1 1 Note that, for purposes of comparison, terms such as “ports” and “fluid spaces” are used to refer to certain prior-art features even if they are different in design and function from equivalent features discussed in the Detailed Description of the present invention.
While all the prior art and this invention latch the engine valve in the closed position by applying the system high pressure P_H and low pressure P_L to the second and first fluid spaces, respectively, resulting in a negative differential pressure that overpowers the net spring returning force, important differences exist in release mechanisms and methodologies.
FIG. 21a depicts the key design and function feature of U.S. Pat. Nos. 4,930,464, 5,595,148, and 5,765,515, which release the piston or engine valve by momentarily reducing the pressure in the second fluid space from P_H to P_L, while keeping the pressure in the first fluid space at P_L. FIG. 21b depicts the key design and function feature of U.S. Pat. Nos. 5,248,123, 5,809,950, 6,167,853, 6,491,007, and 6,601,552, which release the piston or engine valve by increasing the pressure in the first fluid space from P_L to P_H while keeping the pressure in the second fluid space at P_H.
In theory, for both groups in FIGS. 21a and 21b, a net zero differential pressure exists, allowing the spring force complete a pendulum motion. In reality, however, due to fluid inertia and flow losses, the pressures in the first and second fluid spaces are lower and higher than the pressures in the first and second ports, respectively, resulting in a generally negative differential pressure. This greatly impedes engine valve opening and likely stalls the pendulum motion when considering additional engine cylinder air pressure and mechanical frictions on the engine valve. In the case illustrated in FIG. 21a, the first fluid space pressure may easily dip into a negative territory, resulting in cavitations because the piston tries to pull fluid from the first port at the low pressure P_L to the first fluid space.